Roller which incorporates means for moving the roller axially

ABSTRACT

A mechanism for insertion into one end of a distribution cylinder in printing machines for converting the rapid rotational movement of the cylinder to a slow reciprocating cylinder movement. The other end of the distribution cylinder is journalled in a stationary shaft fixedly mounted in a printing machine frame. The mechanism is mounted above the stationary shaft and includes a cylinder (2) which is intended to be fixed in the distribution cylinder. The cylinder (2) is journalled for rotation around the stationary hollow shaft (1). A camming groove element (8; 8A) is fixed axially on the stationary hollow shaft (1) and is journalled for rotation about the symmetry axis (24, 25) thereof. A runner-camming groove-unit (13, 9) facilitates the slow, relative rotation of the camming groove element to an axial reciprocating movement of the cylinder (2), therewith causing the distribution cylinder to move axially backwards-and-forwards.

BACKGROUND OF THE INVENTION

1. Field of the Invention

The present invention relates to distribution cylinders in printingmachines, and more specifically to a distribution cylinder whichincorporates a mechanism which enables the cylinder to move axially,backwards and forwards, at the same time as it rotates.

2. Description of the Related Art

Distribution cylinders are used in printing machines to smooth-out theink layer on one or more printing contra-rotating rollers. By enablingthe distribution cylinder to be driven by the contra-rotating printingroller or rollers at the same time as the cylinder is movedreciprocatingly in the direction of its long axis, the ink layer whichultimately meets the printing platen is smoothed-out or equalized. Poorequalization of the ink layer will result in print defects, such asstriped print.

This axial, reciprocating movement of the distribution cylinder shouldbe a uniform, sinusoidal movement whose frequency is coupled to theprinting speed. This frequency depends on many machine factors, butnormally often lies in the range of 0.5-2 Hz at normal printing speeds.The distribution cylinder should not vibrate at right angles to thecylinder surface, since such vibrations are liable to result inundesirable patterns in the ink layer to be equalized.

This axial movement of the cylinder is typically achieved withconstructions that include levers, reduction gears and camming curves,all of which are fitted externally on the machine frame, withinprotective panels. It is also normal for each inking device to includeup to four distribution cylinders. It will be understood that manymechanical mechanisms of the aforesaid kind are normally required togenerate reciprocatory movement of all of such cylinders.

The incorporation of a mechanism within a distribution cylinder in orderto achieve this axial movement is an old concept which has been appliedin sheet-offset-printing machines. The distribution cylinders in thesemachines do not rotate as fast as the cylinders in modernweb-offset-printing machines. Consequently, it is not necessary for thereduction of the ratio between cylinder speed and the frequency of theaxial movement of the cylinder to be as great. A typical reduction inthese known constructions is 9:1.

A known construction of this kind cannot be transferred toweb-offset-printing machines, since the rollers of such machines rotateat high speeds and the low reduction ratio would then result in an axialmovement frequency of such high magnitude as to risk the occurrence ofharmful vibrations.

It is true that a satisfactory reduction ratio, which preferably lies inthe range 30:1-40:1, could be achieved by incorporating multi-stepgearboxes. Such a construction will afford a number of advantages, suchas high efficiency, long useful life, long periods between servicing,low price, ease of exchange of the whole of the mechanism or partsthereof without disturbing the distribution cylinder, generally speakingindependent of cylinder length.

A construction of this kind, however, also has serious drawbacks, suchas:

a) Strict balancing requirements with regard to the completedistribution cylinder. In the case of cylinder diameters of about 75 mm,the normal maximum imbalance demand is about 6 gcm (gram-centimeter).

b) Low gear mechanism efficiency. Heat emission along the cylinder,which influences the viscosity of the ink and therewith the quality ofthe resultant print.

c) The load-carrying parts of the gear mechanism will have a shortuseful life, due to the large number of moveable parts and the play thatoccurs in time, coupled with relatively frequent services.

d) The construction is also expensive, due to the large number ofmoveable parts.

The German published specification No. 2 045 717 describes adistribution cylinder mechanism which comprises a single-step reductiongear and a cam-curve unit. The reduction gear is comprised of aneccentrically journalled gearwheel which meshes with an internallytoothed annulus connected to the rotary cylinder. With the gearreduction possibilities available at that time, it was possible toachieve a maximum reduction of about 9:1 in this single step. Theexternally-toothed wheel journalled on the stationary eccentrictransmits a slightly increased speed to the cam-curve unit, through themedium of an x-y-link mechanism.

The known distribution cylinder mechanism has two basic features whichrender it unsuitable for use in rapid, web-offset-printing machines,namely:

1) The reduction is too low. This means that axial movement of thecylinder will take place with an impermissibly high frequency, far abovethe desired frequency which, as before mentioned, lies in the range of0.5-2 Hz. In order to obtain sufficiently high reduction, another gearsolution is required, for instance the solution described in my U.S.Pat. No. 5,030,184.

2) The x-y-linkage transfer mechanism is intended to transmit the rotarymotion of the eccentrically journalled gearwheel to the rotational axleof the cam-curve unit. The mass of the x-y-linkage mechanism creates animbalance, resulting in vibrations and frictional heat.

It is possible, of course, to substitute an x-y-linkage system with auniversal drive shaft, diaphragm couplings or arcuate toothed couplings.This would make the construction more complicated, however, andtherewith expensive. In addition, a construction of this kind wouldinclude many components which may become loose because of excessive playand therewith give rise to imbalance and vibrations.

SUMMARY OF THE INVENTION

The object of the present invention is to avoid the aforesaid problems.This object is achieved with an arrangement of the kind defined in theclaims and having the characteristic features set forth therein.

BRIEF DESCRIPTION OF THE DRAWINGS

The invention will now be described in more detail with reference tovarious embodiments of the invention and with reference to the accompanydrawings, in which

FIG. 1 is an axial sectional view of a distribution cylinder providedwith the inventive arrangement;

FIG. 2 illustrates a modified embodiment of the arrangement shown inFIG. 1;

FIG. 3 is an axially sectioned view of still another embodiment of theinventive arrangement;

FIG. 4 illustrates a variant of the arrangement shown in FIG. 3; and

FIG. 5 illustrates another variant of the arrangement shown in FIG. 3.

FIG. 6 illustrates another variant of the arrangement shown in FIG. 3.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

FIG. 1 illustrates an arrangement according to the invention. Thearrangement is contained in a module which is intended to be insertedinto one end of a distribution cylinder, or ink-smoothing cylinder, andfixed thereto. The inventive arrangement is mounted on a hollow shaft.Although not shown, a central distribution cylinder axle extends throughthe hollow shaft 1 and the other end of the distribution cylinder isjournalled on the central axle. This central axle (not shown) isconnected to a printing machine frame. A cylinder 2 is mounted forrotation on the hollow shaft by means of end-walls 3, 4 andneedle-bearings 5, 6. The hollow shaft 1 and the central axle (notshown) are stationary. The distribution cylinder, and therewith thecylinder 2, are driven at a high rotational speed with the aid of meansnot shown. This rotary movement shall be converted to a slow, axialmovement of the cylinder 2 with the aid of the inventive arrangement.The frequency of this axial movement shall be in the order of 0.5 Hz.

In order to move the cylinder 2 axially, the cylinder has a toothed ringor annulus 7 provided with internal teeth on the internal surface of thecylinder. A cylindrical camming toothed element 8 has a camming groove 9at one end thereof and a toothed ring or annulus 10 with external teethon the other end thereof. The camming groove 9 has two camming surfaces11, 12. A roller or runner 13 runs in the camming groove 9 and includesa ball bearing having a cambered or crowned running surface 14. Therunner is provided with a pin 15 which passes through one part 16 of acollar 17 on the end-wall 4. The pin 15, and therewith the runner 13,are fastened to the cylinder 2 by means of a screw 18. The runnerrotates around a rotational axis 19. When seen in the circumferentialdirection of the camming gear element 8, the camming groove 9 has asinusoidal configuration.

The cylindrical camming gear element 8 is positioned obliquely at anangle V degrees in relation to the hollow shaft 1, by means of twobushings 20, 21, mounted on the hollow shaft 1. The camming gear elementis rotatably journalled by means of ball bearings 22, 23 mounted on arespective end of said element. The ball bearing 22 is mounted on thebushing 20 and the ball bearing 23 on the bushing 21.

The symmetry axis of the hollow shaft 1 is referenced 24 while thesymmetry axis of the cylindrical cam gear element is referenced 25. Inone preferred embodiment, the angle V between the symmetry axes 24, 25is 0.45 degrees. The outer cylindrical surfaces of the bushings 20, 21are also at an angle of V degrees in relation to the symmetry axis 24.The bushing 20 has the form of an eccentric annulus. The eccentricity ofthe annulus is chosen so that the external teeth on the toothed annulus10 will mesh with the teeth of the annulus 7. The angled eccentricbushing 20, the ball bearing 22, the toothed annulus 10 and the toothedannulus 7 together form an eccentric gear assembly. The eccentric gearassembly is preferably constructed in the manner described in my U.S.Pat. No. 5,030,184, meaning that the difference between the number ofteeth on the annulus 7 and the number of teeth on the annulus 10 is inthe order of 1 to 2. Since the gear-camming element 8 is inclined, it isalso appropriate to give the annulus 10 a conical configuration with acone angle 2×V degrees. This will result in line abutment between themutually meshing teeth. The teeth on the toothed annulus 7 have an axiallength such as to always achieve meshing engagement between the annulus7 and the annulus 10, irrespective of the axial position of the cylinder2.

The entire assembly comprising camming gear element 8, ball bearings 22,23 and bushings 20-21, is held clamped axially by means of a nut 26,which is preferably screwed very tightly and thereafter fixated with theaid of glue or some corresponding means. The axial position of theassembly on the hollow shaft 1 is fixed with the aid of circlips 27, 28.The angled bushing 21 and the eccentric bushing 20 are affixed by meansof respective cylindrical pins 29, 30, so that the mutual angularposition between said bushings is maintained.

In the case of the preferred embodiment, the toothed annulus 7 hasseventy teeth and the toothed annulus 10 sixty-eight teeth. When thecylinder 2 has completed one revolution, the cylindrical camming gearelement 8 will have rotated one revolution plus two tooth divisions. Inother words, the cylinder 2 and the camming gear element 8 have rotatedtwo tooth divisions in relation to one another. The camming gear element8 thus rotates slowly in relation to the cylinder 2 on which the runner13 is mounted. However, both the cylinder 2 and the camming gear element8 rotate at a very high speed relative to the stationary hollow shaft 1.

During this slow, relative rotation between the cylinder 2 and thecamming gear element 8, the runner 13 moves along the camming groove 9therewith causing the cylinder 2 to move slowly in the direction of itsmain axis. In the preferred embodiment, it is necessary for the cylinder2 to rotate thirty-four (34) revolutions in order to achieve anaxially-reciprocating movement of, e.g., 20 mm top-to-top value. Therunner 13 has an external diameter which is about 0.03 mm smaller thanthe width of the camming groove 9. The runner rolls alternately againstthe one and the other camming surface 11, 12, depending on the directionin which the cylinder 2 moves axially.

In the axial section view of FIG. 1, the bushing 20 is positioned sothat its eccentricity is maximum at its upper defining surface. Thus,when the camming gear element 8 takes the position shown in FIG. 1, thecamming groove 9 will be inclined at an angle of V degrees in relationto the rotational axis 19 of the runner. When the camming gear elementis rotated 90° from this position, the rotational axis 19 and thecamming groove 9 will be parallel. When the camming gear element is thenrotated through a further 90°, the angle defined by the rotational axis19 and the camming surface will be V degrees in the opposite direction(in relation to the position shown in the Figure).

The camming surfaces 11, 12 thus "wobble" through an angle ±V degrees inrelation to the runner rotational axis 19. This "wobbling" movementtakes place at a high frequency and corresponds to the rotational speedof the cylinder 2. A rotational speed in the order of 1200-2000 r.p.m.is not unusual, corresponding to a "wobbling" frequency in the order of20-33 Hz. If the running surface of the runner 13 were to becylindrical, this "wobbling" movement would cause the camming surfaces11, 12 to be clamped against the upper and lower runner edgesrespectively. Such edge abutment is undesirable, since this wouldprevent the runner 13 from rotating, with subsequent damage to therunner ball bearings. One advantage afforded by the present invention isthat the running surface of the runner 13 is cambered (arched). Thiscamber takes-up the "wobbling" movement of the camming surface. As aresult of the camber, the contact between runner and camming surfaces11, 12 is punctiform and the runner runs up-and-down in relation to theequatorial plane of the camber.

Because the camming gear element 8 is positioned obliquely in relationto the symmetry axis 24, the runner will roll on the camming surfaces11, 12 at different radial distances from the symmetry axis 24 of thehollow shaft. This has no deleterious effect, since the runner iscambered and the camber will enable the point of contact to be displacedup and down along the cambered surface.

A ball bearing always has a given degree of self-adjustment, and thisself-adjustment of the ball bearing of the runner 13 further ensuresthat edge abutment will not occur.

The aforedescribed embodiment of the invention can be modified. Onealternative is to provide the camming groove 9 on the inner surface ofthe cylinder 2, and to fix the runner 13 on the camming gear element 8.Another alternative is to choose other ratios between cylinder speed andthe frequency of the axial movement, and also to choose amplitudes andmovement patterns other than sinusoidal. Instead of using a runner inthe form of a ball bearing having a cambered surface, there can be useda spherical bearing.

FIG. 2 illustrates another embodiment of the arrangement according toFIG. 1, in which the bushing 21 and the ball bearing 23 have beenreplaced with a spherical slide bearing 31 which is located centrallybeneath the camming curve as shown in FIG. 2. In this case, the ballbearing 22 transmits the axial movement in both directions, because itsrespective outer and inner rings are fixed at the camming gear element 8and at the bushing 20 by means of locking rings 50, 51. A furtherlocking ring 52 fixes the bushing 20 in the other load direction.

It should be noted that the bushing 21 cannot be given the form of aneccentric bushing having an eccentricity which corresponds to theeccentricity of the bushing 20, in which case the symmetry axis of thecamming gear element 8 would be parallel with and displaced parallel tothe symmetry axis 24 of the hollow shaft 1. The runner 13 would thenroll at varying radial distances from the symmetry axis 24 as it rollsin the camming groove 9, and thus obtain a pulsating movement which issuperposed on the axial, linear movement. This pulsation is extremelytroublesome and cannot be permitted in the case of a distributioncylinder.

In the case of the FIG. 1 and FIG. 2 embodiments, it has been found thatthe runner is subjected to an undesirable acceleration boost as it runsalong the steepest part of the rising parts of the sinusoidal curve. Inthis region of the camming curve, the runner moves "uphill". Thisacceleration boost is expressed as an impact force on the runner,causing the cylinder 2 to be displaced axially through a distancecorresponding to 74 μm. This is an imperfection, or shortcoming, whichis unimportant at low cylinder rotational speeds but which at highcylinder speeds is disadvantageous, because the camming groove willbecome worn in this region of the groove and because the accelerationboost becomes greater with higher cylinder rotational speeds.

In order to eliminate this imperfection, the camming gear element 8 isdivided into two units, viz a camming element 8A and a toothed element8B. The aforedescribed bushing 21 and ball bearing 23 are omitted fromthis embodiment and are, instead, replaced with the ball bearing 23Awhich is fitted directly on the hollow shaft 1. The camming element 8Ais now journalled excentrically around the hollow shaft 1 with the aidof a needle bearing 40 and the ball bearing 23A. Thus, in thisembodiment, the camming surfaces 10 and 11 will always be perpendicularto the symmetry axis 24 during rotation of the camming element 8A. Thisavoids the aforesaid problem of augmented acceleration in the steep partof the camming curve.

The symmetry axis 25A of the cylindrical toothed element 8B is nowinclined at a greater angle to the symmetry axis 24 than in the earliercase. In this case, the angle V is 0.85 degrees. The angle ofinclination is greater, because the eccentricity of the bushing 20 isthe same as in the FIG. 1 embodiment. In view of the high balancingrequirements which prevail at the aforesaid high rotational speeds, theleft end-part 41 of the toothed element 8B is fitted loosely over theneedle bearing 40 and is supported mechanically thereby. When thetoothed element 8B rotates, the end-part 41 will not roll-off on theouter annulus of the needle bearing, but will slide axially on theneedle bearing to some slight extent. The toothed element 8B rotatesabout a stationary symmetry axis 25A which forms an angle 2 V degrees inrelation to the symmetry axis 24 of the hollow shaft, and thisrotational movement is converted to a rotational movement which iscentered around the symmetry axis 24, with the aid of a coupling elementdescribed in more detail herebelow.

According to a first embodiment of the invention, the aforesaid couplingelement is comprised of a number of axially-directed spring pins 42 anda slightly elastic plate 43 which is fitted between the opposingend-surfaces of the camming element 8A and the toothed element 8B. Thespring pins 42 are evenly spaced around the circumferential surface ofthe cylindrical toothed element and are directed axially. The pins 42are pressed into the bore 45 in the end-surface of the camming gearelement 8A and extend freely in a widened part 44 of the bore 45, thepins lying in the bottom part of the bore with a light running fit. Thepins 42 and the plate 43 thus form a coupling which will transmit trueangular movement. One preferred embodiment of the invention compriseseight such spring pins. These spring pins thus transmit the torquederiving from the toothed element 8B.

The axially acting load from the aforedescribed assembly, comprising thebushing 20, the ball bearings 22, 23A, the camming gear element 8A, theplate 43, the toothed element 8B and the coupling, is taken-up by theball bearings 22, 23A. Although the plate 43 is not an imperative partof the coupling element, it affords a given degree of damping axially inthe transmission, which is favorable to the length of useful life of theaxially clamped ball bearings 22, 23A. If the plate 43 is excluded, theopposing end-surfaces on elements 8A and 8B press directly against oneanother. Because of the aforesaid inclination, a gap will always occurbetween the plate and the end-surface of the toothed element 8B, asshown in FIG. 3. This gap will always have the same position in relationto the stationary hollow shaft 1.

The runner 13 of this embodiment of the invention also has a camberedrunning surface 14. If the runner were not cambered, the upper part ofthe runner would strive to rotate at a faster speed than the lower partof said runner, seen in the directions shown in FIG. 3, since the upperpart of the runner is radially spaced from the symmetry axis 24 at agreater distance than the bottom runner part. Slipping would thus occur.

FIG. 4 illustrates another variant of the coupling illustrated in FIG.3, in which the plate and the spring pins have been replaced with avulcanized elastic annulus 46. The annulus is vulcanized in the mutuallyopposing end-surfaces of the elements 8A and 8B.

FIG. 5 illustrates yet another embodiment of a coupling between thecamming element 8A and the toothed element 8B. The coupling illustratedin FIG. 5 is comprised of a disc 47 and a coil spring 48 mounted on theouter surfaces of elements 8A and 8B. The coil spring has two end-parts,of which one is secured in the camming element 8A and the other issecured in the toothed element 8B, as shown at the bottom of FIG. 5.

FIG. 6 illustrates another embodiment of a coupling between the cammingelement 8A and the toothed element 8B. The coupling of this embodimentis comprised of a disc 49A provided with teeth 53, 54 which fit intogrooves 52, 55 provided in the end-surfaces of the camming element 8Aand the toothed element 8B.

Arcuate toothed couplings may also be used instead of the illustratedcouplings. An arcuate toothed coupling is a known element comprised of asleeve having a toothed annulus comprising outstanding teeth which meshwith the internal teeth of a further toothed annulus in a furthersleeve. The sleeves are inserted one into the other so that the teethwill be in engagement with one another, enabling angular transmission ofthe rotary movement.

Other types of homokinetic couplings may also be used.

The invention solves the problems mentioned in the introduction as aresult of the following advantages and fundamental features:

a) Each component can be balanced individually. Very few elements areused. No element is present which can give rise to vibrations andimbalances due to wear, such vibrations and imbalances being likely tooccur when, e.g, x-y-linkage guides, arcuate toothed couplings and thelike are used.

b) The cylindrical camming gear element 8 includes both eccentricannuluses 10 and camming groove. Because the aforesaid gearwheel 10 isslightly conical, good abutment is obtained with the internal tooth andtherewith small losses. The runner 13 is self-adjusting on the cammingsurfaces, thereby avoiding edge abutment. This results in only smalllosses. The cylindrical camming gear element 8 is journalled in ballbearings which are only lightly clamped in an axial direction, resultingin only small losses.

c) Because power losses are small and the temperatures generated inoperation are low, the useful length of life is long, as are also theperiods between servicing.

d) The cylindrical camming gear element 8 replaces many expensive andsensitive components, thereby contributing to a low price.

Although the invention has been described with reference to distributioncylinders, it can be applied equally as well to other types of cylindersor rollers with which rotary movement of the roller shall be convertedto a slow, axially reciprocating roller movement.

I claim:
 1. An arrangement for converting the rotational movement of a roller to an axially reciprocating roller movement, comprisinga stationary shaft (1) having a symmetry axis 24; a cylinder (2) which is to be fixed to said roller, and which is journalled by journals for rapid rotation around the stationary shaft and for axial movement along said stationary shaft; an eccentric gear (7, 10) for reducing the rotational speed of the cylinder, said eccentric gear comprising a toothed annulus (7) having internal teeth and an eccentrically journalled toothed annulus (10) having external teeth; a cylindrical camming gear element (8; 8A) which is fixed axially on the stationary shaft and is journalled by journals obliquely in relation to the symmetry axis (24) at an angle V for rotation around the symmetry axis 24 of said shaft, said camming gear element having at one end thereof the toothed annulus 10 and at the other end a camming groove (9); a coupling means (8; 42, 43; 46; 48; 49A) for transmitting rotational movement of the eccentrically journalled toothed annulus to the camming gear element (8; 8A) and; a runner-camming groove-unit (13) which is mounted in the camming groove (9) and between the camming gear element (8; 8A) and the cylinder (2) for converting rotational movement of the camming gear element to an axially reciprocating movement of the cylinder (2).
 2. An arrangement according to claim 1, charaterized in thatthe cylindrical camming gear element is divided into a cylindrical camming element (8A) with the camming groove (9), and a cylindrical toothed element (8B) with the eccentrically toothed annulus (10); in that the cylindrical camming element (8A) is journalled for rotation around the symmetry axis (24) of the stationary shaft with the aid of journals (40) mounted at one end-surface of the camming element and a bearing (23A) mounted at the other end surface of said camming element; and in that a torque transmission means (42, 43; 46; 48; 49A) connects the camming element (8A) with the toothed element (8B).
 3. An arrangement according to claim 2, charaterized in thatthe torque transmission means includes spring pins (42) which are mounted in axial bores (45) in opposing end-surfaces of the camming and toothed elements (8A, 8B) so as to take-up torque.
 4. An arrangement according to claim 3, characterized by an elastic disc (43) mounted between the opposing end-surfaces of the camming and toothed elements (8A, 8B) so as to provide limited axial springiness.
 5. An arrangement according to claim 2, characterized in thatthe torque transmission means includes a coil spring (48) which connects together the two opposing camming and toothed elements (8A, 8B); and in that the coil spring is anchored at one end in the camming element (8A) and at the other end in the toothed element (8B).
 6. An arrangement according to claim 5, characterized by an elastic disc (47) mounted between the opposing end-surfaces of the camming and toothed elements (8A, 8B) so as to provide limited axial springiness.
 7. An arrangement according to claim 2, characterized in thatthe torque transmission means includes an elastic disc (46) which is vulcanized to each of the two opposing end-surfaces of the camming and toothed elements (8A, 8B).
 8. An arrangement according to claim 2, characterized in thatthe torque transmission means includes a disc having axially outwardly-projecting projections (53, 54) and projection-receiving grooves (52, 55) disposed in the opposing end-surfaces of the camming and toothed elements (8A, 8B).
 9. An arrangement according to claim 1, characterized in thatthe other end of the cylindrical camming gear element (8) is journalled by said journals, said journals including an angled first bushing (21) having a ball bearing (23), and the one end of said camming gear element is journalled by said journals, said journals including an angled second bushing (20) having a ball bearing (22), said second bushing having the form of an eccentric; and in that the angle at which the bushings (20, 21) are inclined is equal to the angle (V) at which the symmetry axis (25) of the camming gear element is inclined to the symmetry axis (24) of the stationary shaft.
 10. An arrangement according to claim 1, characterized in thatthe teeth of the toothed annulus (10) in relation to the axial direction of the cylindrical camming gear element (8) are conical and have an angle of conicity of 2× said V. 